# Shell And Tube Heat Exchanger Design 1.5.4 Crack + License Key/Patch [Updated]

In the case of heat transfer in metal exchangers the thermal resistance of the exchanger wall may be neglected without introducing any appreciable error. Hence, it is permissible to write 14 Differentiating and introducing the Nusselt number in place of the quantities and gives 15 Figure 3. The total power loss is the sum of the losses for the hot and cold mediums. We use Equation 9 and put 16 and obtain for the total energy loss. The change in loss with change of velocity is then given by 17 The change in heat quantity transferred with change of specific loading kF is given by Equation Inserting Equations 12 and 17 in Equation 10 and taking into account Equation 15 gives 18 The coefficients of dNu1 and dNu2 must be equal 0.

This gives two new equations namely, 19a 19b Dividing Equation 19 b by Equation 19 a , inserting for P1 and P2 the values given by Equation 16 and substituting Nu for a1 and a2 gives 20 This equation determines the ratio of the velocities in the right heat exchanger; it does not, however, say anything about the absolute value of the velocities.

It means that heat exchangers in which this ratio of the velocities is observed have for a given surface and a given heat quantity the lowest friction loss. If we take the roots of Equations 19 a and 19 b , and remembering that we obtain as the second condition for the right heat exchanger: Two Equations 20 and 21 completely determine the most favourable heat exchanger.

The resulting transcenddental equation must be solved by trial. The surfaces are found with the aid of the numbers Nu, and Nu, and since Nu is a function of the Reynolds number, the velocities w1 and w2 are also determinate. The dimensions of the heat exchanger are, therefore, fixed.

The Economic Dimensioning of a Heat Exchanger It was seen in the first part that there is a function Z which serves as a criterion of the merit of tube arrangements in heat exchangers. In the second part the conditions fix the dimensions of the surface and sections of the right heat exchanger. The exchanger should, however, like every other apparatus be correctly dimensioned from the economic point of view, that is the total sum of the capital charges and of the running costs should be a minimum.

If P denotes the capital cost, n the interest and deprecition rate, then the capital chargers are. But according to Equation 9 , the energy proportional to F ; hence substituting for differentiating and dividing by P 23 Dividing Equation 22 by Equation 23 24 The total costs in an operational year are a minimum when the capital charges amount to m-times the power costs.

The starting point for this study has been the as sumption of a fixed tube diameter and tube pitch. These and the choice between staggered or straight arrangement of the tubes are determined by dirt deposit and cleaning considerations. How closely these assumptions and the results of the calculation of the right heat exchanger may be adhered to in practice depends on manufacturing conditions, but in any case the above exposition serves as a guide to show in what direction and to what extent modifications are desirable.

The density and heat capacity of both fluids are obtained from tables to standard atmosphere conditions. The characteristic of tube and of the wins are shown in Table 2. Diagram inlets from primary and secondary of the compact heat exchanger are show in Figure 4. Unit thermal water consumption Table 1. Input data of the compact heat exchanger.

Table 2. Characteristic of tube and of the wing. Figure 4. Diagram inlets from primary and secondary fluids. The density of air to Tm is show in Table 3. Table 3. Density of air to Tm. Themal flow Temperature of exit of the water Velocities are: These values are shown in Table 5. Shell-side fluid: Type AES, in. ID Tube bundle: OD, 16 BWG, radial low-fin tubes, 19 fins per inch, ft long, on 1-in.

Heat-transfer area: Sealing strips: Admiralty brass tubes, naval brass tubesheets, all other components of plain carbon steel.

Now we consider modifying the initial design by reducing the number of tubes. If the tube length and overall heat-transfer coefficient remain constant, the number of tubes required is: Reducing the number of tubes and the shell ID will cause both h i and ho to increase, so it may be possible to reduce the tube length as well.

It will also cause the pressure drops to increase. To estimate the effect on tube-side pressure drop, note that from Equation 5. To finalize the design, the overall heat-transfer coefficient must be recalculated, the required tube length determined, and the pressure drops checked, as was done for the second trial in Example 5.

The calculations are left as an exercise for the reader. It can handle all of the TEMA shell types with either plain or radial low-fin tubes. Both single-phase and twophase flows are accommodated on either side of the exchanger. Therefore, this module is used for condensers, vaporizers, and reboilers as well as single-phase exchangers.

Both un-baffled and baffled exchangers are accepted, including the no-tubes-in window configuration. It is important to balance between accuracy and computational cost in order to ensure the most accurate solution in a timely manner. For the purpose of this example, we will define a modified physics induced extremely coarse mesh in order to obtain a relatively quickly solution while still remaining accurate.

In the mesh settings window, change the element size to extremely coarse, then right-click Mesh 1 and select edit physics induced sequence. Click on Free tetrahedral 1, then under the scale geometry section set the x-direction scale field to 0.

Then under the Boundary layer 1 node select Boundary layer properties 1 and define the number of boundaries as 3, then click build all. Once the mesh is built, you are ready to compute Study 1. Please note that it takes about 3 hours to compute the model with an average computer containing 12 Gigabytes of free memory.

Depending on the computational power of your machine, this time may vary. Once the results are calculated, expand the Wall resolution node and click on Surface 1. In the expression section, click replace expression then choose Non-isothermal flow, Upside, and select Wall lift off. Click plot to see the upside wall lift-off for the tubes. From this graph you can see where the most critical areas in terms of mesh resolution are located.

Now that we have validated the model, let us analyze the temperature distribution along all of the wall boundaries. Under the Data Sets node click Surface 1.

From the selection list, choose Walls as the boundary. Under results expand Temperature and click on Surface 1, then change the default temperature unit to degC. Click plot to view the temperature distribution along the new surface. In order to create a 3D streamline view illustrating the velocity and the temperature, we will mirror the solution obtained by the original half heat exchanger geometry. Right-click on data sets, go to more data sets and add a Mirror 3D. From the plane list, choose zx-planes as the plane of symmetry.

Right-click on results and add a 3D Plot group. From the data set list, choose Mirror 3D1. Right-click on 3D Plot Group 4 and add a streamline. In the points edit field under the streamline positioning section, type in to define the number of streamlines in the plot.

Then define the line type as tube in order to make the coloring of the velocity field more visible. Right-click on Streamline 1 and choose Color Expression. From the color table list choose Thermal. Right-click on 3D Plot Group 4 and rename it as Velocity, streamline. This figure illustrates the air flow as well as the water flow through the heat exchanger.

The largely different heat capacities between the air and the water are clearly portrayed in this plot as the temperature of the air changes much more drastically than that of the water when it flows through the heat exchanger.

Now, right-click on Derived Values, and add a Global evaluation. Using the integration and average coupling operators defined earlier, type in the following expression to define the heat transfer coefficient.

Here you can see the global expressions embedded in the terms for: Right-click on Global Evaluations 1 and rename it as Heat Transfer coefficient. Let us now analyze the average water inlet pressure as well as the average air inlet pressure. Right-click derived values and go to average then select Surface average.

## Features of Shell And Tube Heat Exchanger Design 1.5.4

Open in new tab 2. Boundary conditions: Numerical method: There existed fluid maldistribution not only in the longitudinal flow zone of the shell, but also, a more serious fluid maldistribution in the area of the inlet and the outlet.